A toroidal continuously variable transmission has been known which has a toroidal speed change unit that comprises an input disk, an output disk arranged opposite the input disk, and power rollers in frictional contact wit these disks. This toroidal continuously variable transmission continuously changes the rotation speed of the input disk by changing a tilt angle of the power rollers while transmitting the rotation of the input disk to the output disk.
Among the toroidal continuously variable transmissions there are those having a plurality of toroidal speed change units mounted on the same shaft. A type of toroidal continuously variable transmission particularly with two parallelly arranged toroidal speed change units is known as a double cavity type toroidal continuously variable transmission. FIGS. 6 and 7 show one example of the conventional double cavity type toroidal continuously variable transmission. FIG. 6 is a schematic diagram of the conventional double cavity type toroidal continuously variable transmission and FIG. 7 is a cross section showing a transmission ratio control mechanism in the double cavity type toroidal continuously variable transmission of FIG. 6 that includes the toroidal speed change units and a transmission ratio control valve. FIG. 8 is a flow chart showing the procedure for determining a duty of a solenoid valve in the transmission ratio control mechanism of FIG. 7.
In the double cavity type toroidal continuously variable transmission shown in FIG. 6, two toroidal speed change units 1, 2 are mounted side by side on a main shaft 3. The toroidal speed change unit 1 comprises an input disk 4, an output disk 5 disposed opposite the input disk 4, and power rollers 6 in frictional engagement with toroidal surfaces of the disks 4, 5. The toroidal speed change unit 2, like the toroidal speed change unit 1, comprises an input disk 7, an output disk 8 disposed opposite the input disk 7, and power rollers 9 in frictional engagement with toroidal surfaces of the disks 7, 8. The toroidal speed change units 1, 2 each have two sets of power rollers 6, 9. The power rollers 6, 9 each can be rotated on their own rotation axis 10 and tilted about a tilt axis 11 (perpendicular to the sheet of the drawing) crossing the rotation axis 10 at right angles.
In the toroidal speed change units 1, 2, the input disks 4, 7 are displaceable in the axial direction of the main shaft 3 and rotatable together with the main shaft 3. The driving force of the engine is supplied through a torque converter 12 to an input shaft 13 laid on the same axis as the main shaft 3. At the front end of the input shaft 13 there is provided a loading cam 14 with a cam roller 15. The rotation of the loading cam 14 causes the input disk 4 and, through the main shaft 3, the input disk 7 to rotate. Thus, the main shaft 3 functions as an input shaft for the input disks 4, 7. The cam action of the loading cam 14 generates a thrust force (in the axial direction of the main shaft 3) pressing the input disks 4, 7 against the power rollers 6, 9 according to the magnitude of the input torque. The thrust force holds the power rollers 6, 9 between the input disks 4, 7 and the output disks 5, 8 according to the magnitude of the torque being transmitted. The rotation of the input disks 4, 7 is transmitted to the output disks 5, 8 through the power rollers 6, 9 based on the oil shearing force.
In each of the toroidal speed change units 1, 2, the power rollers 6, 9 are tiltable about the tilt axes 11, and the rotation of the input disks 4, 7 while being transmitted to the output disks 5, 8 is continuously changed in speed according to the tilt angle of the power rollers 6, 9. The power rollers 6, 9 are supported rotatable and pivotable relative to trunnions 33, 37 (see FIG. 7) so that they can be pivoted corresponding to their displacement in the axial direction of the main shaft 3 caused by the thrust force.
The output disks 5,8 are coupled back to back on a coupling shaft 16 as by spline so that they can rotate together. The coupling shaft 16 is a hollow shaft mounted relatively rotatable on the main shaft 3. A sprocket 18 is formed integral with an intermediate portion of the hollow shaft. The coupling shaft 16 is connected to an output shaft 26 as described later. The output disks 5, 8 are supported on the casing 19 by bearings (not shown) that support thrust and radial loads through the coupling shaft 16. The driving force transmitted to the output disks 5, 8 is further transferred through a chain transmission device 17 as a first transmission means, which includes the sprocket 18, a chain 20 and an intermediate sprocket 21, and then to a countershaft 22 having the intermediate sprocket 21 at one end thereof.
At the other end of the countershaft 22 is mounted a forward clutch 23, whose output side is connected to a first gear 24 which in turn is in mesh with a second gear 25 secured to the output shaft 26 of the entire transmission. Hence, the forward clutch 23 can switch the countershaft22 and the first gear 24 between an idling state and a torque transmitting state. The first gear 24 and the second gear 25 form a reverse transmission means that reverses the rotation of the countershaft 22 and transmits it to the output shaft 26. The mechanism ranging from the chain transmission device 17 as the first transmission means to the countershaft 22 to the first and second gears 24, 25 as a second transmission means forms a reverse transmission mechanism that reverses the rotation of the output disks 5, 8 and transfers it to the output shaft 26.
Between the main shaft 3 and the output shaft 26 is arranged a planetary gear mechanism 27, which comprises a sun gear 28 secured to the main shaft 3, pinions 30 meshing with the sun gear 28 and having a carrier 29, and a ring gear 31 meshing with the pinions 30 and secured to the output shaft 26. Between the carrier 29 and the casing 19 is installed a reverse clutch 32 that switches the carrier 29 between an idling state and a fixed state with respect to the casing 19.
Next, referring to FIGS. 7 and 8, the speed change control as performed by the toroidal speed change units 1, 2 will be described. The power rollers 6, 9 are rotatably supported on the trunnions 33, 37 by rotary support shafts 34, 38. The trunnions 33, 37 have tilt axes 11 and can be moved in the axial direction of the tilt axes 11 and pivoted about the tilt axes 11. That is, as the power rollers 6, 9 tilt, the tilt angle displacements .theta. of the power rollers 6, 9 result directly in rotary displacements of the trunnions 33, 37 about the tilt axes 11.
The input disks 4, 7 and the output disks 5, 8 are elastically deformed in the axial direction of the main shaft 3 by the thrust. Because the axial positional references of the toroidal speed change units 1, 2 are determined by the casing 19 on which the output disks 5, 8 are supported by the bearings (not shown), the power rollers 6, 9 are displaced in the axial direction of the main shaft 3 according to the elastic deformations of the disks. The rotary support shafts 34, 38 that rotatably support the power rollers 6, 9 on their end portions 36, 40 are offset from pivotable support shafts 35, 39 pivotably supported on the trunnions 33, 37, so that the displacements of the power rollers 6, 9 in the axial direction of the main shaft 3 are absorbed by the precess motions of the power rollers 6, 9 about the pivotable support shafts 35, 39. When the positions in the thrust direction of the output disks 5, 8 are determined with respect to the casing 19, the positions of the power rollers 6, 9 are determined and the positions in the thrust direction of the input disks 4, 7 are also determined.
The structures of hydraulic cylinders 42, 45 to displace the trunnions 33, 37 in the tilt axis direction are basically identical, and their identical constitutional elements are assigned like reference numerals. The tilt axes 11 of the trunnions 33, 37 are provided with pistons 41, 44 respectively. The piston 41 is slidably installed in the hydraulic cylinder 42 formed in the casing 19. The hydraulic cylinder 42 is partitioned by the piston 41 to form therein a speed reduction side cylinder chamber 43A and a speed increasing side cylinder chamber 43B. When a pressure difference occurs between the speed reduction side cylinder chamber 43A and the speed increasing side cylinder chamber 43B in the hydraulic cylinder 42, the trunnions 33, 37 move together with the power rollers 6 in the tilt axis direction. When oil pressure is supplied to the speed increasing side cylinder chamber 43B, the speed is changed toward the speed increasing side. When oil pressure is supplied to the speed reduction side cylinder chamber 43A, the speed is changed toward the speed reduction side. Oil paths 47A, 47B in the toroidal speed change unit 2, as in the toroidal speed change unit 1, communicate to the hydraulic cylinders 45 each having the corresponding speed reduction side cylinder chamber 46A, speed increasing side cylinder chamber 46B and piston 44.
In a body (valve case) of a spool valve 48 a sleeve 49 is slidably installed. First springs 50 engaging the ends of the sleeve 49 urge the sleeve 49 to be held at a neutral position. Slidably installed inside the sleeve 49 is a spool 51 which is urged toward right in FIG. 7 by a second spring 52 disposed at one end of the spool 51. The other end of the spool 51 is engaged by a precess cam 53 through a lever 54. The spool valve 48 has an SA port formed at one end and an SB port at the other end, with the SA port supplied with a control oil pressure PA through a solenoid valve 55A and with the SB port supplied with a control oil pressure PB through a solenoid valve 55B. The spool valve 48 also has a PL port connected to a line pressure (oil pressure source), an A port communicating to the speed reduction side cylinder chamber 43A through the oil path 47A, a B port communicating to the speed increasing side cylinder chamber 43B through the oil path 47B, and an R port communicating to a reservoir.
This toroidal continuously variable transmission has a variety of sensors, such as a car speed sensor 56 and an accelerator depression sensor 57 for detecting the amount of depression of the accelerator pedal. Speed change information such as car speed v and accelerator depression amount Acc detected by these sensors is supplied to a controller 58. The controller 58 outputs to the solenoid valves 55A, 55B control signals corresponding to a target transmission ratio calculated based on these speed change information. The solenoid valves 55A, 55B have an output port C for outputting the control oil pressure PA, PB to the SA or SB port of the spool valve 48, a drain port D, and an oil pressure source port E communicating with the control oil pressure source (pilot oil pressure source) PS. The solenoid valves 55A, 55B are of the same type and, in the case shown, are normally open changeover valves whose output port C communicates with the oil pressure source port E when an electric failure such as wire break occurs with the solenoid valves. The solenoid valves 55A, 55B, upon receiving the control signals from the controller 58, either supply the oil pressure from the control oil pressure source PS to the SA port and SB port or release the control oil pressure from the SA port and SB port to the drain to move the sleeve 49 in the axial direction according to the target transmission ratio. The sleeve 49 is formed with communication holes corresponding to the ports PL, R, A and B and, according to the position of the spool 51, allows the ports PL and R to communicate with the port A or port B.
The front end of the tilt axis 11 of one of the trunnions 33 in the toroidal speed change unit 1 is connected to a precess cam 53. The lever 54 pivotably supported at its central part has one end thereof engaged with the precess cam 53 and the other end engaged with the other end of the spool 51 of the spool valve 48. The precess cam 53 detects a combined displacement of the tilt axis direction displacement Y and the tilt angle displacement .theta. of the trunnion 33. The spool 51 of the spool valve 48 is moved according to this combined displacement. The spool valve 48 and the solenoid valves 55A, 55B together form a transmission ratio control valve that receives a control signal representing a target transmission ratio from the controller 58 and a signal representing the combined displacement from the precess cam 53 and control the oil pressure of the hydraulic cylinders 42, 45.
At the neutral position where the tilt axis direction displacement Y is zero, the trunnions 33, 37 maintain the tilt angle of the power rollers 6, 9 as it is at this moment and hold the transmission ratio at a value present at that time. That is, at this neutral position, the trunnions 33, 37 are located at a position in the tilt axis direction where the rotation axes 10 of the power rollers 6, 9 cross the rotating center line of the input disks 4, 7 and output disks 5, 8. At this position, the power rollers 6, 9 are tilted through a tilt angle displacement corresponding to that transmission ratio. The spool 51 is moved following the sleeve 49 that was shifted according to the target transmission ratio and assumes the position where it closes the A port and B port.
The solenoid valves 55A, 55B may be duty solenoid valves that can change a time rate of the operated position that the valve disk can take by changing the duty ratio of a pulse current used to energize the electromagnetic coils. One example procedure of determining the duties of the solenoid valves 55A, 55B is shown in FIG. 8. According to the duty determination procedure shown in FIG. 8, the car speed v and the accelerator pedal depression amount Acc are detected by the car speed sensor 56 and the accelerator depression sensor 57 (in step 20, abbreviated S20 hereafter), and the controller 58 calculates the duties for energizing the solenoids of the solenoid valves 55A, 55B according to the target transmission ratio that was determined based on the speed change information such as car speed v and accelerator pedal depression amount Acc (S21). That is, if we let the duties for the solenoid valves 55A, 55B be duty A and duty B, respectively, they can be expressed as
duty A=f (Acc, v) (f represents a function) PA1 duty B=100%-duty A
Because the solenoid valves 55A, 55B are normally open solenoid valves of the same type, the duty A and the duty B that make up the total duty (100%) are distributed so that a pressure difference is normally produced between the SA port and the SB port. The duties thus obtained are output to the corresponding solenoid valves 55A, 55B (S22) before the program returns to the main routine.
Because the solenoid valves 55A, 55B are energized under duty A and duty B, respectively, the sleeve 49 moves until the pressure difference between the control oil pressures PA and PB supplied from the control oil pressure source PS through the solenoid valves 55A, 55B to the ports SB and SA at the ends of the spool valve 48 balances with the force of the first springs 50. The position to which the sleeve 49 has moved represents the target transmission ratio.
The transmission ratio is changed by displacing the trunnions 33, 37 from the neutral position in the axial direction of the tilt axes 11. That is, when during the rotation of the input and output disks a target transmission ratio is changed and the sleeve 49 is shifted to a different position, a relative movement occurs between the sleeve 49 and the spool 51. Depending on the amount of the relative displacement, the A port or B port communicates with the PL port connected to the line pressure, displacing the trunnions 33, 37 in the tilt axis direction. As the trunnions 33, 37 are displaced in the tilt axis direction, the power rollers 6, 9 are also displaced in the tilt axis direction, which in turn causes the positions of contact between the power rollers 6, 9 and the input disks 4, 7 and output disks 5, 8 to move from those contact positions associated with the trunnion's neutral position. As a result, the power rollers 6, 9 receive a tilt force from both disks and begin to tilt about the tilt axes 11 in a direction and at a speed determined by the direction (Y&gt;0 or Y&lt;0) along the tilt axes 11 and the amount (absolute value of Y) of the displacement. As the contact points between the both disks and the power rollers change, the speed change is performed continuously.
The combined displacement of the tilt axis direction displacement Y and the tilt angle displacement .theta. of the trunnions 33, 37 as detected by the precess cam 53 acts on the other end of the spool 51 of the spool valve 48 through the lever 54 to move the spool 51 against the force of the second spring 52 acting on the one end of the spool 51. Hence, depending on the relation between the position of the sleeve 49 given as a target transmission ratio and the position of the spool 51 given by the precess cam 53, the A port and B port connected to the oil paths 47A, 47B are switched between the PL port and the R port to connect the oil paths 47A, 47B to the PL port or R port to produce a pressure difference between a working oil pressure Pdown in the oil path 47A and a working oil pressure Pup in the oil path 47B. The pressure difference between the working oil pressures of the cylinder chambers 43A and 43B causes the trunnions 33, 37 in the toroidal speed change units 1, 2 to move in the tilt axis direction, starting the speed change operation. As the tilt angle approaches a target tilt angle and the tilt axis direction displacement Y of each trunnion 33, 37 approaches zero, the position of the spool 51 given by the precess cam 53 nears the position of the sleeve 49 corresponding to the target transmission ratio until their positional difference progressively converges to zero, at which time the speed change operation is terminated.
As described above, the tilt angle control of the power rollers 6, 9 is performed so that the spool 51 will follow the sleeve 49 that occupies the position corresponding to the target transmission ratio set by the controller 58. Until the sleeve-following action by the spool 51 is completed, the working oil pressure supplied to the hydraulic cylinders 42, 45 through the open ports is controlled to displace the trunnions 33, 37 in the tilt axis direction. When the tilt axis direction displacement Y of the trunnions 33, 37 is zero and the power rollers 6, 9 are tilted through the target tilt angle, the sleeve-following action by the spool 51 is completed.
There is also a toroidal continuously variable transmission in which the trunnions supporting the power rollers are not displaced even when the power rollers receive a tangential force from the input disk and the output disk during torque transmission (see Japanese Patent Laid-Open No. 151218/1995). This toroidal continuously variable transmission produces a pressure difference between two cylinder chambers-used to control the tilt axis direction displacement of the trunnions-that cancels the tangential force acting on the power rollers to maintain the trunnions at a neutral position and thereby prevent the restarting of the speed change operation after its completion.
The above-described prior arts have the following drawbacks. That is, in the conventional toroidal continuously variable transmissions shown in FIGS. 6 and 7 as examples, the solenoid valves 55A, 55B that control the control oil pressures PA, PB acting on the both ends of the sleeve 49 are normally open duty solenoid valves, a type in which the solenoid valves 55A, 55B produce the pressure of the oil pressure source PS at their output ports C when there is no control signal from the controller 58. It should be noted that the solenoid valves 55A, 55B are not completely free from electric failures and the most likely electric failure is a wire break.
When an electric failure occurs with the solenoid valve 55A, the control oil pressure PA that appears at the SA port of the spool valve 48 is the pressure of the control oil pressure source PS and becomes larger than the control oil pressure PB at the SB port (PA.gtoreq.PB). Hence, the sleeve 49 is pushed toward the right in FIG. 7 and the working oil pressures to the hydraulic cylinders 42, 45 have the relation of Pup.gtoreq.Pdown, causing the trunnions to move toward the speed increasing side. When on the other hand an electric failure occurs with the solenoid valve 55B, the control oil pressure PB at the SB port is the pressure of the control oil pressure source PS and becomes larger than the control oil pressure PA at the SA port (PA.ltoreq.PB). Hence, the working oil pressures to the hydraulic cylinders 42, 45 have the relation of Pup.ltoreq.Pdown, causing the trunnions to move toward the speed reduction side. When both the solenoid valves 55A, 55B should fail, the SA port and the SB port of the sleeve 49 both produce the pressure of the control oil pressure source PS, fixing the sleeve 49 at the center position and setting the transmission ratio at an intermediate between the maximum speed decrease ratio and the maximum speed increase ratio. Of these electrical failures, the wire break in the solenoid valve 55B causes the toroidal continuously variable transmission to make a rapid shiftdown, resulting in possible engine overruns and vehicle spins, giving rise to a dangerous condition.
As described above, there are problems that need to be solved: when the solenoid valves should electrically fail, a rapid shiftdown of the toroidal continuously variable transmission must be avoided by preventing the speed change operation of the toroidal speed change units from being performed toward the speed reduction side; and a measure to prevent a rapid shiftdown in the event of electric failures, when implemented, must not affect the normal transmission ratio control and must ensure the speed change control is performed stably.
It is an object of this invention to provide a toroidal continuously variable transmission that solves the above problems and in which even when one or both of two solenoid valves that control the position of a sleeve constituting a transmission ratio control valve are failed electrically, the control oil pressures output from the electrically failed solenoid valves to the spool valve are set to other than a pressure that causes the trunnions to move in the tilt axis direction toward the speed reduction side, thereby avoiding a rapid shiftdown of the transmission and preventing occurrence of dangerous conditions such as engine overruns and vehicle spins. It is also an object of this invention to provide a toroidal continuously variable transmission that can prevent a measure, adopted to prevent a rapid shiftdown in the event of such electrical failures, from affecting the normal transmission ratio control and thereby stably maintain the transmission ratio.